The field of gas bearings has long been considered theoretically ideal for large ultrahigh speed machinery. However, it has been disappointing in practice because the load carrying capacity of bearings of this nature, which theoretically should be great, has been found in practice to be disappointingly small. I have ascertained some of the causes for this disappointing performance and avoided them with a gas hydrodynamic bearing capable of supporting heavier loads at high speed.
Hydrodynamic fluid bearings are essentially of two varieties: (a) hard surface types, and (b) compliant surface types. The hard surface types include (1) plane surface; (2) contoured surface types, such as spiral groove, tapered-land, step, pocket; and (3) tilt pad. Hydrodynamic supporting fluid films are generated over hard surface types by dragging the lubricating fluid, whether it be oil, water or air, by viscous shear forces, into a converging space between the two bearing members. The shape of that space is determined, in the case of plane and contoured surface types, by the profile to which the bearing surfaces are machined, and also by the position and orientation of the bearing surface of the rotating member. The tilt pad type employs a pad mounted on a pivot which enables the tilt pad to tilt under the influence of hydrodynamic forces acting on it to assume a slope that will cause a supporting hydrodynamic fluid film to be generated.
The compliant surface type of bearing employs a compliant support layer which supports a flexible bearing sheet such as a thin sheet of stainless steel. Under the influence of relative movement between the opposing bearing surface and the bearing sheet, hydrodynamic forces are generated which depress the compliant support and the overlying bearing sheet to a profile that is conducive for generating a supporting hydrodynamic fluid film between the opposing bearing surface and the bearing sheet.
The hydrodynamic supporting fluid film is created by the viscous or shear forces acting in the fluid parallel to the direction of relative movement between the two bearing surfaces. A rotating thrust runner, for example, drags its boundary layer of air with it as it rotates opposite to a tilt pad. The boundary layer, in turn, drags in the immediately adjacent layer of air, and so forth. In this way, an air velocity gradient is established in the gap between the thrust runner and the tilt pad. The pad is supported at about 58% of its length in the direction of rotation of the thrust runner and tilts away from the thrust runner at its leading edge. This creates a wedge-shaped gap between the thrust runner and the pad which causes the fluid being dragged into the wedge to increase in pressure toward the trailing edge of the tilt pad. This pressure increases gradually to a maximum at approximately 3/4 of the pad length from the leading edge.
In the same way, compliant surface bearings also create a hydrodynamic supporting fluid wedge by viscous drag of the boundary layer exerted by the relative movement between the rotor bearing surface and the compliant bearing surface. A zone of high pressure fluid is created which provides the fluid support to maintain the separation between the opposing bearing surfaces.
A pressure curve showing the distribution of pressure over a hard, tilt pad bearing surface reveals that the pressure increases gradually from the leading edge of the wedge shaped gap between the opposing bearing surfaces to a maximum at about 3/4 of the length, then falls off steeply. This pressure curve inherently results in a peak. Although the maximum pressure in the supporting fluid film might be high, the total supporting force is not high because of the nonuniform pressure distribution over the surface of the bearing. The pressure curve over a compliant bearing module reaches an early maximum pressure which is maintained over a substantial portion of the module surface area before falling off steeply at the end of the module. The total supporting force over the compliant bearing module is higher than that over the rigid tilt pad because the area under the compliant bearings's pressure curve is considerably greater than the area under the tilt pad's pressure curve, even though the maximum pressure over the compliant bearing may be less.
One of the advantages of gas bearings is their ability to operate in high temperature environments. However, it is common for a thrust bearing to operate in a very uneven temperature condition, with the thrust runner being hotter at its bearing face than on the opposite face, establishing a temperature gradient across the bearing member in an axial direction. The higher temperature zone at the bearing face causes the material of the bearing to expand nonuniformly and produce a convex shape of the bearing face. This thermal distortion of the bearing surface has an adverse effect in the hydrodynamic action at the bearing interface which no longer has the optimum wedge for generating the hydrodynamic supporting fluid film. In effect, the distortion reduces the bearing surface and transfers the entire load to the portion of the bearing surface which has not been bowed away from the opposed bearing surface by thermal distortion. The smaller effective bearing surface now must carry the same load, and the resulting greater pressure may exceed the load bearing capacity of the bearing. This condition may be further exacerbated by misaligned loads on the thrust bearing. Where the shaft tilts or precesses under the misaligned load, one edge of the thrust runner lifts away from the thrust plate and the entire axial load must then be borne by the opposite edge.
The damping effect of hydrodynamic fluid bearings is theoretically ideal for ultrahigh speed applications. However, the runout excursions which should be damped by the bearing, may be of a nature which do not coincide with the bearing's damping characteristics. Thrust runner excursions exerted on a thrust bearing can vary widely in frequency amplitude and direction. In addition, fluid effects, such as half speed whirl, can exist in the lubricating fluid itself to complicate the situation. Failure of a bearing can occur when the power of the rotor runout excursions exceeds the damping capacity and misalignment tolerance of the bearing, such as when a rotor, passing through its critical velocity, experiences runout of such amplitude that the hydrodynamic fluid film is breached and the bearing surfaces contact each other at high relative speed.
Many prior art hydrodynamic bearings have been unsuccessful because of the lack of understanding of what occurs in the bearing interface and how the complex forces acting on the bearing interact and affect the hydrodynamic action of the bearing. When solutions are proposed, they are usually focused on a single, perceived problem in the bearing and fail to account for the other effects. I have designed a bearing which accounts for the factors which are most influential in the operation of the hydrodynamic fluid bearing and which solve the problems which most seriously affect its load bearing capacity.